GB2468656A

GB2468656A - Friction clutch that modulates the clamp load in relation to rotational speed

Info

Publication number: GB2468656A
Application number: GB0904512A
Authority: GB
Inventors: Grahame Smith; Jonathan Grant
Original assignee: AP Racing Ltd
Current assignee: AP Racing Ltd
Priority date: 2009-03-17
Filing date: 2009-03-17
Publication date: 2010-09-22
Anticipated expiration: 2029-03-17
Also published as: GB2468656B; GB0904512D0

Abstract

The invention relates to a friction clutch assembly adapted to automatically modulate the force with which the friction plates are clamped (the clamp load) during engagement of the clutch in response to changes in the rotational speed of the clutch housing. The clutch assembly (10, fig 1) comprises a counter-pressure member 20, a clutch housing 12 rotationally fast with the counter-pressure member 20, a pressure plate 16 rotationally fast with the housing 12 but movable in an axial direction of the clutch relative to the housing 12, at least one fiction plate 14, 15 between the pressure plate 16 and the counter-pressure member 20 and a diaphragm clutch spring 18 for clamping the friction plate(s) between the pressure plate 16 and the counter-pressure member 20. The clutch housing 12 is configured to deform radially as a function of its rotational speed, resulting in a change in the axial spacing (a, a') between the axial position 34 at which the clutch spring 18 reacts with the housing 12 and the counter-pressure member 20. The change in axial spacing (a, a') alters the cone angle of the diaphragm spring 18 and varies the proportion of the spring force applied to the pressure plate 16 during clutch engagement. Material having a higher density then the housing as weights (38, fig 5 42, fig 6 or 46, fig 7) may be used to radially deform the housing.

Description

Friction Clutch This invention relates to friction clutches for motor vehicles. In particular, but not exclusively, this invention relates to friction clutches for use in cars, more particularly high performance andlor racing cars.

In a typical motor vehicle friction clutch, one or more friction plates are positioned between a pressure plate and a counter-pressure member or reaction plate. The counter-pressure member will often be a flywheel but this is not always the case. The counter-pressure member and pressure plate are often in driving connection with an output shaft of the engine and rotate about a common axis. One or more of the friction plates will be in driving connection with an output shaft of the clutch, which may comprise an input shaft of an associated gearbox of the vehicle. Such clutches are typically engaged by means of a diaphragm spring that acts between the clutch housing and the pressure plate. The diaphragm spring biases the pressure plate towards the counter-pressure member in order to clamp the friction plates between pressure plate and the counter-pressure member.

The force clamping the friction plates is referred to as the clamp load" and the magnitude of the clamp load is a factor in determining the amount of torque that can be transmitted through the clutch. Generally speaking, the higher the clamp load the more torque can be transmitted through a clutch of any given size, The actual clamp load applied to the friction plates when the clutch is engaged or partially engaged depends on the force applied to the pressure plate by the diaphragm spring less any forces acting in the opposite direction. In some known clutches, the pressure plate is attached to the housing by means of spring straps which bias the pressure plate away from the flywheel to ensure a clean disengagement. In clutches of this type the clamp load will be substantially equal to the force applied to the pressure plate by the diaphragm spring less the force of the spring straps which act in the opposite direction.

However, in other clutch arrangements there are no significant forces acting on the pressure plate in the opposite direction to the spring and so that the clamp load in these cases will be substantially equal to the force applied to the pressure plate by the spring. When a clutch is only partially engaged, part of the spring force is taken by the release bearing and only a proportion of the available spring force is applied to the pressure plate to clamp the friction plates.

Diaphragm springs are conical and the spring force produced by a diaphragm spring usually varies in a non-linear maimer in dependence on its cone angle. This is known as the characteristic curve of the spring. Many clutches are configured so that the cone angle of the spring when the clutch is engaged and the friction plates are new is such that the spring force is just below its peak value. As the friction plates wear, the cone angle of the spring when the clutch is engaged increases leading initially to an increase in the spring force until it passes its peak, after which further increases in the cone angle as the plates continue to wear results in a reduction of spring force.

The types of friction clutches used in racing cars are such that pedal travel between a clutch release position and a clutch engaged position is very small. This in effect means that the clutch engagement is sudden and sharp.

Effective control of clutch engagement is particularly important for a racing driver to ensure that power is transmitted to the driving wheels in a manner that enables them to pull away from the start line as quickly as possible. Sudden engagement of the clutch may result in excessive wheel spin or even stalling of the engine. Typically, a driver will wish to hold the clutch in a partially engaged condition during which time only part of the available engine torque is transmitted through the clutch, until the vehicle has gathered sufficient speed and the clutch can be fully engaged. The position at which the clutch is held is sometimes referred to as the "bite point".

Holding a racing clutch on the bite point is difficult as only a relatively small amount of pedal travel can lead to a significant variation in the clamp load and hence the torque transmitted through the clutch. This problem is exacerbated were the clutch incorporates so called carbon/carbon composite clutch plates, i.e. clutch plates formed from a carbon based matrix filled with a carbon based filer. It is a property of carbon/carbon composite clutch plates that their coefficient of friction increases significantly as the temperature of the plates increases, until the plates reach a saturation temperature beyond which the coefficient of friction remains largely static or may even decrease slightly. It is a further property of carbon/carbon clutch plates that they expand as they heat up.

When a friction clutch is held in a partially engaged condition, the friction plates are not fully clamped and so slip relative to one another. As the plates rub against one another their temperature will increase. Where some or all of the friction plates are made of a carbon/carbon composite material, this increase in temperature may lead to an increase in the coefficient of friction of those plates and hence an increase in the amount of torque transmitted through the clutch, even if the clutch pedal is held stationary. At the same time, the carbon/carbon composite friction plates expand as their temperature rises. As the plates expand they become more firmly clamped between the pressure plate and the flywheel and cause the clamp load to increase, again without any change in the clutch pedal position. Changes in the coefficient of friction and expansion of the carbon/carbon friction plates happens very quickly and a driver may have insufficient time to react by depressing the clutch pedal in order to compensate for the increase in torque being transmitted through the clutch. This may lead to excessive wheel spin at the start of the race, to the engine being stalled if the engine speed is too low or even to a false start.

Alternatively, if the driver depresses the pedal too far in an attempt to compensate for the increasing clamp load, the clamp load may fall leading to a reduction in the torque transmitted through the clutch. This will typically result in an increase in the engine speed as the load on the engine is reduced.

It is an objective of the present invention to provide an improved friction clutch which overcomes, or at least mitigates, some or all of the disadvantages

of the prior art clutches.

It is also an objective of the invention to provide an improved friction clutch in which engagement can be more easily controlled than in prior art clutches.

In accordance with a first aspect of the invention, there is provided a friction clutch assembly comprising a counter-pressure member, a clutch housing rotationally fast with the counter-pressure member, a pressure plate rotationally fast with the housing but movable in an axial direction of the clutch relative to the housing, at least one friction plate between the pressure plate and the counter-pressure member and a clutch spring operative between the housing or a component mounted to the housing and the pressure plate to urge the pressure plate towards the counter-pressure member so as to clamp the at least one friction plate between the pressure plate and the counter-pressure member, in which the clutch assembly comprises a mechanism for modulating the clamp load in dependence on the rotational speed of the clutch housing, at least during engagement of the clutch.

The counter pressure member may be an integral part of the housing or it may be a separate component such as a flywheel.

In one embodiment, the clutch housing is configured to deform as a function of its rotational speed so as to vary the axial spacing between the axial position at which the spring reacts with the housing or component and the counter-pressure member.

An axially extending region of the housing may be configured deform radially outwardly as the rotational speed of the clutch housing increases, resulting in a reduction of the axial spacing between the axial position at which the clutch spring reacts with the housing or component and the counter-pressure member.

Weights of a material having a higher density than the material(s) of the clutch housing may be mounted in or to the axially extending region of the clutch housing.

The clutch assembly may be a multi-plate clutch assembly having a plurality of first friction plates rotationally connected with the clutch housing and a plurality of second friction plates rotationally connected with an output hub arrangement.

The clutch housing may comprise a radially extending plate member, a clutch cover, and a plurality of circumferentially spaced axial drive lugs extending between the plate member and the clutch cover, the drive lugs being configured to deform radially in dependence on the rotational speed of the housing. In which case, at least one weight may be mounted in or to each of the drive lugs. In this embodiment, either the clutch cover or the radially extending plate member may comprise the counter-pressure member.

The clutch spring may be one or more diaphragm springs operative between a first fulcrum on or mounted to the housing and a second fulcrum on the pressure plate. In which case, the clutch may be configured such that radial deformation of the housing varies the axial spacing between the first fulcrum and a portion of the counter-pressure member against which the at least one friction plate is clamped. The clutch assembly may be configured such that in use, radial deformation of the housing varies the cone angle of the, or each, diaphragm spring, in dependence on the rotational speed of the clutch housing.

In embodiments where the clutch spring is a diaphragm spring, diaphragm spring may have a plurality of release levers and the assembly may have a release bearing movable for engagement with the release levers to engage and disengage the clutch, in which case, the clutch may be configured such that in use when the clutch is partially engaged, deformation of the housing in dependence on the rotational speed of the clutch housing varies the relative proportions of the spring force taken by the pressure plate and the release bearing.

In accordance with a second aspect of the invention, there is provided a motor vehicle comprising a friction clutch in accordance with the first aspect, an engine and a gearbox, in which the clutch housing is rotationally fast with an output shaft of the engine and at least one friction plate is drivingly connected with an input shaft of the gearbox. The arrangement may be configured such that in use, the torque transferred across the clutch is modulated as a function of engine speed, at least when the clutch is partially engaged.

Several embodiments of the invention will now be described, by way of example only, with reference to the accompanying drawings, in which: Figure 1 is perspective view of a first embodiment of a clutch assembly in accordance with the invention at rest; Figure 2 is a view similar to that of Figure 1 but showing the clutch assembly at when rotating at maximum RPM; Figure 3 is a cross sectional view through the clutch assembly of Figure 1; Figure 4 is a composite diagrammatic view illustrating the effects of radial deformation of the housing of the clutch assembly of Figures 1 and 2, the upper half shows a cross sectional view of part of the clutch assembly in a rest condition whilst the lower half shows a cross sectional view of the same part of the clutch assembly when rotating at maximum RPM; Figures 5, 6 and 7 are views similar to that of Figure 3, but showing the clutch assembly when rotating at maximum speed and illustrating alternative embodiments of the clutch housing in accordance with the invention; and, Figure 8 is a graph showing axial deflection against rotational speed for a typical example of a clutch housing for use in a clutch assembly in accordance with the invention and for an equivalent conventional clutch housing.

With reference initially to Figures 1 to 4, a clutch assembly 10 in accordance with the present invention comprises a clutch housing 12 within which is located a stack of first and second friction plates 14, 15, a pressure plate 16 and a clutch spring 18.

The clutch housing 12 in this embodiment is of the so called "basket" type and comprises a radially extending annular plate member 20, a plurality of circumferentially spaced drive members 22 extending axially from an outer diameter region of the plate member 20, and a clutch cover 24 secured to the free ends of the drive members 22. The annular plate member 20 extends from a central flange 26 which is configured for driving engagement with an output shaft of an engine (not shown), so that the clutch housing 12 rotates together with the output shaft about a conmton axis Y. The stack of first and second friction plates 14, 15 are located within the housing 12. The first plates 14 (which will be referred herein as drive plates) have radially outwardly extending lugs 28 on their outer diameter which locate between the axial drive members 22 so that they are constrained to rotate with the housing about the axis Y but are moveable relative to the housing in the axial direction. The second plates 15 (which will be referred to as driven plates) are interleaved between the drive plates 14 and each driven plate 15 is mounted to a central hub 30 so that the hub is constrained to rotate with the driven plates and via versa. The hub 30 has internal splines 32 for mounting on an output shaft (not shown) that may be an input shaft of an associated gearbox (also not shown). The driven plates 15 are arranged so as to be movable in an axial direction of the clutch relative to the clutch housing 12.

The pressure plate 16 is located between the stack of friction plates 14, and the clutch cover 24. The pressure plate 16 has radially extending teeth 1 6a on its outer diameter for engagement with the drive members 22 so that the pressure plate 16 is constrained to rotate with the housing about the axis Y whilst being able to move in the direction of the axis Y relative to the housing.

The clutch spring 18 comprises a diaphragm spring which acts between a fulcrum 34 on the clutch cover and a fulcrum 36 on the pressure plate 16.

When the clutch is engaged, as shown in Figure 3, the spring 18 biases the pressure plate 16 towards the radially extending plate member 20 so as to clamp the friction plates 14, 15 between the pressure plate 16 and a face 20a of the radially extending annular plate member 20. The face 20a is located at a radially outward position on the annular plate member 20 and constitutes the counter-pressure member. In this embodiment, the diaphragm spring 18 is a single diaphragm 18 but the spring could comprise two or more spring members in a known manner.

The clutch assembly 10 is a so-called "pull-type" clutch in which a radially outer region of the diaphragm spring 18 contacts the fulcrum 34 on the clutch housing whist the spring 18 contacts the fulcrum 36 on the pressure plate 16 at a position radially inward of the fulcrum 34 on the clutch housing. The spring 18 has a number of radially inwardly projecting release leavers 35 that are pulled in the direction of arrow A by a release bearing 37 (illustrated schematically in Figure 4) to release or disengage the clutch. However, the invention can be equally applied to so-called "push-type" clutches.

In the present embodiment, there are no bias forces opposing the force applied to the pressure plate 16 by the diaphragm spring so that when the clutch is engaged and the whole of the available spring force is applied to the pressure plate, the clamp load will be substantially equal to the available spring force, subject to any losses in the system due to friction etcetera. When the clutch is only partially engaged, the release bearing 37 engages with the radially inner ends 35a of the release levers to hold the spring in a partially engaged potion so that only a proportion of the available spring force is applied to the pressure plate 16. The cone angle of the spring 18 when the clutch is partially engaged will be different from that when the clutch is engaged so that the available spring force will also vary. Whether the change in the available spring force is an increase or a decrease will depend at the relative positions of the spring along its characteristic curve when the clutch is engaged and partially engaged.

The clutch assembly 10 is arranged so that the clamp load is automatically modulated during clutch engagement in dependence on the rotational speed of the engine to which the clutch housing 12 is mounted. In the present embodiment, this is achieved by configuring the clutch housing 12 so that it deforms as a function of its rotational speed in order to vary the minimum axial spacing between the fulcrum 34 on the clutch housing and the face 20a of the radially extending annular plate member 20 against which the friction plates 14, 15 are clamped. This in turn alters the cone angle of the diaphragm spring and varies the clamp load. This effect will now be described in more detail with reference to Figure 4.

The top half of Figure 4 shows the clutch assembly 10 at rest and with the clutch partially engaged. In this condition the axial distance between the fulcrum 34 on the housing and the face 20a is "a" and the cone angle of the diaphragm spring 18 is "0". The release bearing 37 is in contact with the radially inner ends 35a of the release finger to hold the clutch in a partially engaged condition so that a proportion of the available spring force is taken by the release bearing 37.

The bottom half of Figure 4 shows the clutch assembly 10 at maximum RPM in the same partially engaged condition. As shown, the axial drive members 22 have deformed radially outwardly at the centre and the radially inner end of the clutch cover is deflected axially towards the engine side of the clutch (to the right as shown), relative to an axially fixed datum such as the central flange 26. The radially outer region of the annular plate member 20 is also deflected axially in the same direction as the clutch cover relative to the datum 26 but by a lesser extent. The net effect of this is to displace fulcrum 34 on the clutch cover axially towards the face 20a so that the axial distance "a" between the fulcrum 34 and the face 20a when the clutch is rotating at maximum speed is reduced when compared to the distance "a" when the clutch is at rest.

Axial displacement of the fulcrum 34 towards the face 20a reduces the axial spacing between the fulcrum 34 on the cover and the fulcrum 36 on the pressure plate and thus alters the cone angle 0' of the diaphragm spring. In the present embodiment, the cone angle 0' is reduced when the clutch is rotating at maximum RPM when compared to the cone angle 0 of the diaphragm spring 18 when the clutch is at rest. In alternative embodiments in which the diaphragm spring is mounted with a smaller cone angle initially, the change in the deformation of the housing when the clutch is rotating may cause the diaphragm spring to go flat or indeed to reverse the cone angle.

Due to the characteristic curve of the diaphragm spring 18 in this embodiment, the reduction in cone angle 0' as the housing deforms results in a reduction in the available spring force produced by the diaphragm spring 18.

However, a more significant effect of the axial displacement of the fulcrum 34 and the change in cone angle is that the radially inner ends 35a of the release fingers are deflected axially (to the left as shown in Figure 4) by a distance X so as to unload the release bearing 37. This movement results in a greater proportion of the available spring force being applied to the pressure plate 16 and an increase in the clamp load applied to the friction plates. In Figure 4, the release bearing is shown as being fully unloaded when the clutch housing is rotating at maximum RPM. However, this is for illustration purposes only and the movement of the release fingers may only partially unload the release bearing.

As described above, the annular plate member 20 and the cover 24 are both deflected towards the engine side of the clutch (to the right as shown in Figure 4) as the housing deforms. As a result, the internal contents of the clutch housing, including the friction plates 14/15, the pressure plate 16 and the diaphragm spring 18 are also moved slightly in this direction so that the spring fingers are moved closer to the release bearing. This movement is in the opposite direction to the movement of the release fingers resulting from the change in cone angle of the spring. However, movement of the spring fingers caused by the shift in the internal contents of the clutch housing is less than the movement in the opposite direction caused by the change in cone angle.

Accordingly, the overall effect of the deformation of the housing is to reduce the loading on the release bearing and to increase the force applied by the spring to the pressure plate 16 and consequently to increase in the clamp load.

In alternative clutch configurations, movement of the internal contents of the clutch housing as the housing deforms may move the spring fingers in the same axial direction as the movement resulting from the change in cone angle and so contribute to unloading of the release bearing. It will be understood, therefore, that the designer must take into account the relative movements of the clutch assembly components as the housing deforms and changes in the available spring force to ensure that a desired increase in clamp load is achieved for a given increase in the rotational speed of the clutch housing.

It will be appreciated that if the speed of rotation of the clutch housing 12 is reduced from the maximum RPM to 0, the effects described above will be reversed and the clamp load reduced.

It can be seen then that in a clutch assembly in accordance with the invention, the clamp load applied to the friction plates 14, 15 by the diaphragm spring 18 when the clutch is partially engaged is automatically increased as the speed of rotation of the clutch housing increases and decreased as the speed of rotation of the clutch housing decreases. Since torque transfer through the clutch is a function of the clamp load, automatic modulation of the clamp load can be used to modulate torque transfer whilst the clutch is being engaged and in particular whilst the clutch is held in a partially engaged condition.

For reasons of clarity, Figure 4 illustrates the extreme conditions when the clutch housing is not rotating and when it is rotating at maximum speed. In use when the clutch is held in a partially engaged position, the clutch housing will be rotating at a speed somewhere between these extremes and the variation in rotational speed will be considerably less. It will be appreciated, therefore, that the range of radial deformation of the housing 12 and the changes in the cone angle of the diaphragm spring 18 required to modulate the clamp load will typically be less than that shown in Figure 4.

Operation of the clutch under various conditions will now be discussed.

If the driver of a vehicle to which a clutch assembly 10 in accordance with the invention is fitted wants to hold the clutch on the so called "bite point", the clutch pedal is released to partially engage the clutch until sufficient torque is being transmitted to either hold the vehicle steady or to initiate rolling movement of the vehicle as required and the driver will try to maintain the engine at a steady speed somewhere between the rest and the maximum RPM.

At this stage, the clutch housing 12 is rotating at the same speed as the engine so that the drive members 22 are deformed radially outwardly by an amount between the two extremes shown in Figure 4 and the release bearing is in engagement with the release fingers 35 and so is taking a proportion of the spring force produced by the diaphragm spring.

Should the torque transfer across the clutch suddenly increase as a result of poor clutch pedal control by the driver or due to an increase in the Co..

efficient of friction of the carbon/carbon friction plates andlor to an expansion of the plates as discussed previously, the engine speed will decrease as greater load is placed on it. As the engine speed falls, the rotational speed of the clutch housing will also fall reducing the radial deformation of the clutch housing 12 and increasing the axial spacing between the fulcrum 34 in the clutch cover and the face 20a. Accordingly, the cone angle 0 of the diaphragm spring 18 will be increased and the radially inner ends of the release fingers 35 will move axially (to the right as shown in the drawings) bringing them into closer contact with the release bearing. This loads the bearing, increasing the proportion of the available spring force which is taken by the release bearing so that the clamp load is reduced. The reduction in clamp load compensates, at least partially, for the increase in the co-efficient of friction and/or the expansion of the friction plates so that torque transmitted through the clutch is reduced, bringing it back to, or at least closer to, the level desired by the driver.

Alternatively, if there is a reduction in the torque transmitted through the clutch, due say to slippage between the friction plates or due to the driver inadvertently depressing the clutch pedal, the engine speed will increase. This will give rise to an increase in the rotational speed of the clutch housing leading to an increase in the clamp load and a corresponding increase in the torque transmitted to compensate for the initial reduction.

Modulation of the clamp load and the torque transmitted as discussed above happens automatically in response to changes in the engine speed and much more quickly than a driver could be expected to react.

I

Operation of the clutch 10 has been described above in relation to a situation in which a driver wishes to hold the clutch on the bite point, i.e. at a constant partially engaged condition. However, the ability to modulate the clamp load as function of the engine speed is also beneficial when a driver wishes to gradually engage the clutch over a period of time in order to feed the torque through to the driving wheels in a controlled manner. During this process, modulation of the clamp load as described above will assist the driver in engaging the clutch smoothly despite changes to the coefficient of friction of the plates and/or expansion/contraction of the plates and other factors that may affect the torque transfer across the clutch.

It will be appreciated that clutch engagement is a complex dynamic process and that it is difficult to describe in detail all the changes which may take place as the clutch is engaged. However, in essence the invention resides in the provision of a clutch assembly having a housing which is configured to deform as a function its rotational speed in order to modulate clamp load when the clutch is partially engaged.

Conventional clutch housings are designed to be substantially rigid in use. Whilst some radial deformation of the housing might occur this will be limited and not sufficient to cause a significant change in the clamp load as a function of its rotational speed. Accordingly, in a clutch assembly 10 in accordance with the invention, the clutch housing 12 is designed to be sufficiently flexible to enable the required amount of deformation to take place over the expected range of rotational speeds, whilst ensuring that the housing has sufficient structural strength to function as a clutch housing in its intended application. The design of the drive members 22 in particular may be modified from that used in conventional basket type clutch housings to promote radial deformation. As illustrated in Figure 4, the drive members 22 may be arranged so that they curve inwardly towards the centre of the clutch when the clutch is at rest and to deform so as to curve or bow outwardly away from the centre of the clutch at higher rotational speeds. This increases the range of possible radial deformation when compared to drive members that are generally straight when the clutch is at rest.

Centrifugal weights having a higher density than the material of the housing may be mounted to, or incorporated in, the drive members 22 to promote outward radial deformation of the drive members as the speed of rotation increases. For example, the clutch housing may be made of a comparatively lightweight material such as titanium alloy or aluminium alloy and the centrifugal weights made from a higher density material such as tungsten.

Figures 5, 6, & 7 illustrate different ways in which the centrifugal weights can be mounted, by way of non-limiting example.

In Figure 5, a weight 38 having a relatively small length is mounted in a recess 40 in the inner surface of each drive member at a position approximately mid-way along its length. The weight may be an interference fit in the recess so that it is held in place sufficiently tightly that it does not fall out when the clutch is at rest. When the clutch is rotating, centrifugal force will assist in keeping the weight in position.

In the embodiment shown in Figure 6, one or more weights 42 are located inside an internal bore 44 in each in the drive members. The bore 43 may be an extension of a bore formed to receive a bolt 45 which secures the clutch cover to the drive member 22. As shown in Figure 1, two bolts may be used to secure the clutch cover to each drive member so that each drive member may be provided with two corresponding bores 44 with one or more centrifugal weights in each bore.

In the embodiment shown in Figure 7, a centrifugal weight 46 is again located in a recess 48 on the inner side of each drive member 22. However, in this embodiment, the weight 46 extends over the majority of the length of the drive 22, though it may be shaped so as to have a greater mass in a mid-region of the drive member. As in the embodiment shown in Figure 5, the weight may be an interference fit in the recess so that it is held in place sufficiently tightly that it does not fall out when the clutch is at rest.

The provision of a recess 40, 48 or one or more bores 44 to receive the weight(s) may also assist in making the housing flexible enough to accommodate the required level of deformation.

The amount of axial deflection required to achieve the desired modulation of the clamp load will depend on a number of factors including the size of the clutch housing and its speed of rotation. Figure 8 is a graph illustrating axial deflection of a clutch housing against rotational speed for a typical example of a clutch housing adapted for use in a clutch assembly in accordance with the invention and for an equivalent conventional clutch housing. In the example shown, the clutch housings are for a 100mm diameter clutch and the axial deflection is measured between two datum points, one on the face 20a of the radially extending plate member 20 and one on the inner surface of the clutch cover 24. The axial deflections have has been established using Finite Element analysis configured to simulate a clutch in a partially engaged state and which is subjected to a partial clamp load of 2kN to simulate a partially engaged condition. Full clamp load for the clutches analysed being 16.5 kN. The graph shows a maximum RPM of 18000 and at this speed of rotation it can be seen that the axial deflection (i.e. the reduction in the axial distance between the two datums) for a conventional clutch housing is typically in the range of 0.05 mm to 0.10 mm. In contrast, a clutch housing adapted for use in a clutch assembly in accordance with the invention has an axial deflection in the range 0.20 mm to 0.40 mm. As illustrated by the graph, the axial deflection in a clutch housing for use in a clutch assembly in accordance with the invention is significantly higher than the maximum axial deflection in a conventional equivalent clutch housing at all rotational speeds above 0 RPM.

In particular, it is noted that in a clutch housing in accordance with the invention, the axial deflection at maximum RPM is in excess of 0.15 mm whilst the maximum axial deflection for the equivalent conventional clutch housing is 0.10 mm.

In the embodiments described above, the drive members 22 are formed integrally with the radially extending annular plate member 20 but this is not essential. The drive members 22 could be formed separately from the annular plate member 20 and mounted it using any suitable means. The drive members 22 could, for example, be formed integrally with the clutch cover if desired. In addition, the clutch 10 could be configured so that the friction plates 14/15 are clamped against a face on the clutch cover 24 rather than the radially extending plate member 20. In this arrangement, the clutch cover 24 would become the counter-pressure member and its radial extent increased. The pressure plate 16 and diaphragm spring 18 would then be located on the opposite side of the clutch pack between the friction plates 14/15 and the radially extending plate member 20 with the clutch housing fulcrum 34 on the radially extending plate member 20.

Furthermore, the invention is not limited to application in multi-plate clutches having a basket type housing with axially extending drive members but can also be applied to multi-plate clutches having a generally cylindrical clutch housing with teeth on its inner surface for driving engagement with corresponding teeth on the drive plates. All that is necessary is that the housing is configured so that the axially extending region of the housing deforms radially as function of the speed of rotation of the clutch housing to vary the axial position of the diaphragm spring fulcrum 34 on the clutch housing relative to the counter-pressure member.

The invention can also be applied to friction clutch assemblies having only a single friction plate andlor to clutch assemblies of the type having a housing mounted to a flywheel, which comprises the counter-pressure member.

Furthermore, the principles disclosed can be applied to clutches having clutch springs other than diaphragm springs.

The foregoing embodiments are not intended to limit the scope of protection afforded by the claims, but rather to describe an example as to how the invention may be put into practice. Those skilled in the art will appreciate that the clutches in accordance with the invention can be produced

Claims

Claims 1. A friction clutch assembly comprising a counter-pressure member, a clutch housing rotationally fast with the counter-pressure member, a pressure plate rotationally fast with the housing but movable in an axial direction of the clutch relative to the housing, at least one friction plate between the pressure plate and the counter-pressure member and a clutch spring operative between the housing or a component mounted to the housing and the pressure plate to urge the pressure plate towards the counter-pressure member so as to clamp the at least one friction plate [0 between the pressure plate and the counter-pressure member, in which the clutch assembly comprises a mechanism for modulating the clamp load in dependence on the rotational speed of the clutch housing, at least during engagement of the clutch.
2. A friction clutch assembly as claimed in claim 1, in which the clutch housing is configured to deform as a function of its rotational speed so as to vary the axial spacing between the axial position at which the spring reacts with the housing or component and the counter-pressure member.
3. A friction clutch assembly as claimed in claim 2, in which an axially extending region of the housing is configured deform radially outwardly as the rotational speed of the clutch housing increases, resulting in a reduction of the axial spacing between the axial position at which the clutch spring reacts with the housing or component and the counter-pressure member.
4. A friction clutch assembly as claimed in claim 3, in which weights of a material having a higher density than the material(s) of the clutch housing are mounted in or to the axially extending region of the clutch housing.
5. A friction clutch assembly as claimed in any one of the preceding claims in which the clutch is a multi-plate clutch having a plurality of first friction plates rotationally connected with the clutch housing and a plurality of second friction plates rotationally connected with an output hub arrangement.
6. A friction clutch assembly as claimed in any one of the previous claims, in which the clutch housing comprises a radially extending plate member, a clutch cover, and a plurality of circumferentially spaced axial drive lugs extending between the plate member and the clutch cover, the drive lugs being configured to deform radially in dependence on the rotational speed of the housing.
7. A friction clutch assembly as claimed in claim 6 when dependent on claim 4, in which at least one weight is mounted in or to each of the drive lugs.
8. A friction clutch assembly as claimed in claim 6 or claim 7, in which the clutch cover comprises the counter-pressure member.
9. A friction clutch assembly as claimed in claim 6 or claim 7, in which the radially extending plate member comprises the counter-pressure member.
10. A friction clutch assembly as claimed in any one of the previous claims, in which the clutch spring comprises one or more diaphragm springs operative between a first fulcrum on or mounted to the housing and a second fulcrum on the pressure plate.
11. A friction clutch assembly as claimed in claim 10, the clutch being configured such that radial deformation of the housing varies the axial spacing between the first fulcrum and a portion of the counter-pressure member against which the at least one friction plate is clamped.
12. A friction clutch assembly as claimed in claim 10 or claim 11, in which the clutch is configured such that in use, radial deformation of the housing varies the cone angle of the, or each, diaphragm spring, in dependence on the rotational speed of the clutch housing.
13. A friction clutch assembly as claimed in any one of claims 10 to 12, in which diaphragm spring comprises a plurality of release levers and the assembly further comprises a release bearing movable for engagement with the release levers to engage and disengage the clutch, in which the clutch is configured such that in use when the clutch is partially engaged, deformation of the housing in dependence on the rotational speed of the clutch housing varies the relative proportions of the spring force taken by the pressure plate and the release bearing.
14. A friction clutch assembly substantially as hereinbefore described, with reference to and as illustrated in Figures 1 to 4, or Figure 5, or Figure 6 or Figure 7 of the accompanying drawings.
15. A motor vehicle comprising a friction clutch as claimed in any one of the previous claims, an engine and a gearbox, in which the clutch housing is rotationally fast with an output shaft of the engine and at least one friction plate is drivingly connected with an input shaft of the gearbox.
16. A motor vehicle as claimed in claim 15, in which the arrangement is configured such that in use, the torque transferred across the clutch is modulated as a function of engine speed, at least when the clutch is partially engaged.